High pressure pumping system



C. J. COBERLY HIGH PRESSURE PUMPING SYSTEM March 22, 1960 4 Sheets-Sheet 1 Filed July.l2, 1957 PUMP HEH @HT/O (ENGINE HREH Y L .K E E. me. w M me R A L C BY H/S ATTORNEYS. HAR/els, K/scH, Fos Tel? e; He kms NNUU March 22, 1960 c. J. coBERLY 2,929,327

HIGH PRESSURE PUMPING SYSTEM Filed July 12, '1957 4 Sheets-Smet 2 /Nl/NTOR. CLARE/vee d. CoeR/EY c. J. coBERLY 2,929,327 HIGH PRESSURE PUMPING SYSTEM March 22, 1960 Filed July 12. 1957 4 Sheets-Shea?I 3 PLUNGE R00 DIH/METER PLU/VGE 0IA/METER PUMP AREA RAT/o (ENG/NE HREF) /NVs/VTOI?. CLARE/vee d. Cosi/nv BY H/S H TTORNEYS.

Hn/aR/s, Knee/4, FosTa/e 6C HAeR/s March 22, 1960 c. J. coBERLY 2,929,327

HIGH PRESSURE PUMPING SYSTEM Filed July l2. 1957 4 sheets-sheet 4 Fig. 9.

l. 20 l. l O

ACTUAL OPERATING PRESSURE. HCTUHL, OPERATING PRESSURE THEORET/CHL OPERATING PRESSURE REC/PROGRA. OF NET PRESSURE LOSS /N PERCENT OF HCTUHL. OPERHT/NG PRESSURE PUMP AREA RAT/0 (ENG/NE AREA BY H/S ATTORNEYS. HAR/as, KlecH, Fos-rek 6E HARP/s nited States i tent i 2,929,327 Patented Mar. 22, 1.960

HIGH PRESSURE PUMPING SYSTEM Clarence J. Coberly, San Marino, Calif., assigner to Kobe, gne., Huntington Park, Calif., fa corporation of Caliorina Application July 12, 1957, Serial No. 671,622

7 Claims. (Cl. 10S-46) As general background, any uid operated pump in` cludes a pump section actuable by a fluid operated engine section to pump uid from a well in which the pump is disposed upwardly to the surface through a production tubing set in the well. pump includes cylinder means having an inlet in communication with the well and having an outlet in communication with the lower end of the production tubing. Reciprocable in the cylinder means is an engine and pump piston means, comprising interconnected engine' and pump piston elements, vfor pumping uid from the well upwardly through the production tubing to the: surface, reciprocatory movement of the engine and pump More particularly, such a" piston means resulting from the alternate application of operating uid pressure and spent operating uid or exhaust pressure thereto by an engine valve means forming partof the pump. The engine and pump piston means includes a transverse pump area which supports the en tire column of production iluid between it and the'- surface,

and which moves such;column upwardly through the production tubing, during a working; stroke of the piston means, the pump -being provided with a workingvalve which prevents bacldlow into the Well during such working stroke. In order to produce the working stroke of the piston means, it is provided with an engine area to which the operating iluid pressure is applied at least during the working stroke. In the case of a single-acting pump, to which the present disclosure will be restricted for convenience without necessarily limiting the invention thereto, the working stroke of the piston means occurs in one direction only and movement of the piston means in the opposite direction merely constitutes a return stroke, the pump being provided with a standing valve which prevents backow from the production tubing into the well during such return stroke. Normally, the pump is so oriented and constructed that the working stroke of the engine and pump piston means occurs in the upward direction, the return stroke of the piston means thus being the downward stroke thereof.

A iluid operated pump of the foregoing nature may be utilized in either an open system, wherein the spent operating iluid is mixed with the production fluid and is returned to the surface through the production tubing,

or in a closedsystem wherein the spent operating uid is conveyed to the surface through a separate, return tubing. In the case of an open system, the exhaust pressure acting on the engine section of the uid operated pump is the production'column pressure, while, in a closed system, the exhaust pressure acting thereonisthe return column pressure. For convenience, the present disclosure will be restricted to a closed system without necessarily limiting the invention thereto.

Heretofore, uid operated pumps, such as those disclosed in my Patents Nos. 2,081,220 and 2,473,864, and in my aforementioned copending application, have been provided with engine areas which are large in comparison to the pump areas thereof so that the pumping sys temsin which such pumps are incorporated must utilize large volumes of operating fluid at relatively low operating Huid pressures. For example, conventional lluid operated pumps-are customarily provided with ratios of pump. area to engine area ranging from 0.65 :l to 1.5 :l

with the laverage being approximately 1:1. With such" ratios, the operating fluid pressures utilized range approxifluid circulates are experienced, particularly in wells having high productive capacities. Such fluid friction losses may be reduced to some extent by enlarging the tubings of the uid operated pumping system through which the operating iluid flows, but this is expensive and the extent to whichsuchtubings can be enlarged is severely restricted by the small diameter of the well bore.

Another serious disadvantage of prior uid operated pumps is that the large engine-areas previously utilized severely restrict the sizes of the production uid passages through the .pumps and the sizes of the working and standing valves incorporated therein, the sizes of such L passages and valves conventionally being selected on the basis of a compromise dictated bythe space occupied by the engine section of the pump. As will readily be understood, Ithe small production uid passages and valves resulting from such a comprise produce serious fluid friction losses within the pump itself, these being in addition to the losses in the tubings of the system result-v ing 4from the large-volumes of operating fluid which must be handled in a low pressure system wherein the pumps have low ratios :of pumparea to engine area.

-Anv important object of the present invention is to over- I come the foregoing and various other disadvantages of prior practice by providing a fluid operated pump having a high ratio of pump area to engine area, i.e., a pump area to engine area ratio of at least 3:1, and to utilize a correspondingly high operating fluid pressure. As will appear hereinafter, the operating uid pressure utilized in accordance with the present invention may attain values as high as 20,000 to 25,000 p.s.i. gauge or more, depending on the production rate, the pump depth setting, the specific ratio of pump area to engine area employed, the sizes of the tubings through which the operating iluid flows, and on other factors to be considered.

Numerous advantages and new results are derivable from the high pressure operation, which may logically bereferred to as super pressure operation, attainable as a consequence of providing a iluid operated pump v with a pump area to engine area ratio of 3:1, or higher.

tanfalvamagel Since the' present invention results in substantial reductions in operating fluid friction losses, smaller tubings for the operating uid may be employed, which has the advantage of materially reducing the cost of the system. Super pressure operation also markedly increases the capacity of the pump, or the same capacity'maybe attained with a much smaller unit.

Factors such as the foregoing enable the present invention to reduce drilling and casing costs for a specified pumping capacity as the result of reductions in tubing and pump sizes since a smallerwell bore may be used.

Expressed differently, the present invention results in aminimum capital investment in drilling and casingthe.

well, and in the hydraulic pumping system, per unit ofv production capacity.

.Another feature of the invention is that the quantity` o f operating fluid required per unit of production capacity is minimized so that the system for supplying. the operating fluid under pressure is reduced in size. further reduces the over-all capital investment per unit of production capacity.

Another important advantage of the invention` is thatthe surface lines for conducting operating fluid to and from the well may be materially smaller. This still further reduces capital costs per unit of production capacity,V

particularly where one source of operating fluid serves a number of wells so that the total length may be quite large.

Another advantage of super pressure operation is that,

as will be explained in more detail hereinafter, the pres sure, utilized maintains the viscosity of the operating iluid.

substantially constant throughout the system, Le., the pressure utilized increases the viscosity to an extentsuicient to substantially offset the reductionin viscosity due..

to the` normal increase, in. temperature with well depth. Maintaining the viscosity substantially constant through out the system in this manner limits leakage losses in. the pump and improves lubrication, which are important features.

Another and extremely important advantage of the present invention is that utilizing a high pump areato engine area vratio and a high operating fluid pressure minimizes the size of the engine section of the lluid operated pump so that the size of the pump section there- ThisA of surface lines- ...acaasev 4y to the pump disclosed in my aforementioned copending application. In this pump, the engine and pump piston means comprises a single plunger the upper end of which provides the pump area, the engine area being a small, downwardly-facing annulus which encircles the large axial passage for the production uid. With a ratio of pump area to engine area` of at least 3:1 in accordance with the invention, the size of this downwardly-facing annulus which forms the -enginearea is quite small so `spaced from the plunger.

of islimited substantially only by the over-all dimensions of the pump. The result of this is that the pump section may be' designed solely on the basis of considerations related to eflicient pumping, there being no necessity for: p

a compromise in the design of the pump section resulting from the necessity for allocating a great deal ofthe space available in the iluid operated pump to the engine'section thereof, as inY the case of prior lluid operated pumps.

Thus, the present invention provides a pump section having large passages for the production lluid, large valves, and the like, all of whichA minimize fluid friction losses with respect to production liuid ow through the pump.

The super pressure operation of thepresent inventionattains particular utility when applied to a tiuid operatedpump wherein the engine and pump piston means comprisesa single plunger having a large axial passage there'- through for the production Huid and carrying a large working valve which prevents -flow through such passage during the working stroke of the plunger, the housing of the pump being provided with a correspondingly large standing valve, in either the inlet or the outlet of the pump, which prevents backllow into the Well during the return stroke of the plunger.- By employing a ratio4 of pump area'to engine area of 3:1, or higher, with such a pump, the diameter of the axial, production-Huid passage through the combined engine and pump plunger is a substantial proportion of the over-all diameter of the plunger. As will be apparent, this reduces productionuid friction losses through the pump to a minimum, which is an important advantage.

that the diameter of the axial production-iluid passage through the plunger is a substantial proportion of the over-al1 diameter of the plunger to minimize duid friction losses.

With this pump, the working stroke ofthe plunger is elfected by applying the exhaust pressure to an upwardlyfacing annular areal of the plunger while applying the operating fluid pressure to the downwardly-facing annular engine area mentioned, these two annular areas being located at the ends of a. reduced-diameter section of the plunger.V The return stroke of the plunger is effected in the presently/'preferred embodiment of the invention by applyingthe operating fluid pressure to the upwardlyfacing annular area While still applying thev operating fluid pressure to the downwardly-facing annular area, or by'applying the exhaust. pressure to both of these areas.

`The upwardly-facing annular area is very slightly larger than the downwardly-facing annular area so that the return stroke of the plunger results from the application of the.. operating lluid pressure, or the exhaust pressure,

to both of these annular areas.

Another advantage of utilizing a uid operated pump having. the structure disclosed in my aforementioned copending application in connection lwith a ratio of pump .areaz tov engine area of 3:1. or higher, is that the engine valvemeans, which produces reciprocatory movement of theplunger by alternately applying the exhaust pressure and the operating fluid pressure to the. upwardly-facing annular area of the plunger, is located alongside of and ference with the pump section ofthe lluid operated pump by the engine section thereof so that the pump section design is determined virtually entirely by pump considerationsalone.`

Theforegoing objects, advantages, features and results 'of the present invention, together withl various other objects, advantages, features. and resultsthereof which will` be. evident toV those skilled` inv the lluidoperated pump art in the light of this. disclosure, may be' attained vvithy the exemplary embodimentV of the inventionl describedY in detail'hereinafter and illustrated in the' accompanying'.

drawings, in which: v

Fig. l is a view, partially in Vertical section, of an installation of a uid operated pumping` system of the inand thus results in an increase inthe diameter of an axial While the fluidoperated pump of the invention may l take other forms,v the invention is particularly applicable production-fluid-passage through such plunger; j u

Fig. 8\ isy a graph illustrating the manner in which an increasing ratio ofv pump areavto engine area results in a decrease in the operating uid pressure loss in the system due to friction for different tubingsizes; and

Fig.l 9 is agraph illustrating the effect of anfincreasing. ratio of pump area to engine area o n the reciprocal of the uid pressure loss in percent off theaotual orvtotal operatingnuid pressure.

This further eliminates interaalsmeer Referring rst to Figs. present invention is illustrated therein as installed in a well `provided with a casing having perforations 12 through which production fluid may ow into the casing from the productive formation, not shown, into which the well is drilled. The casing 10 is surmounted by a casing head 14 from which operating fluid supply and return tubings 16 and 18 and a production tubing 20 are suspended. Connected to the lower ends ofthese tubings is a bottomhole pump housing 22.

Hydraulically movable through the production tubing 20 between the surface and an operating position within the housing 22 is a free, fluid operated pump unit 24. As will be discussed in more detail hereinafter, the pump unit is of the reciprocatory type and includes a pump section for pumping uid from the well upwardly through the -production tubing 20 to the surface, the production tubing being large as compared to the supply and return tubings to permit the pumping therethrough of a large volume of production fluid with minimum iiuid friction losses and to permit the use of a pump unit of maximum diameter. The pump section of the pump unit 24 is actuated by a fluid operated engine section. The latter is actuable byan operating fluid under pressure delivered theretothrough the supply tubing 16, the spent operating uid being discharged at exhaust pressure into the return tubing 1-8 which conveys it to the surface. Since, as will be discussed hereinafter, the engine section of the pump unit 24 is actuated with a small volume of operating uid at a very high pressure, the supply and return tubings 16 and 18 are small relative to the production tubing 20. Z

'The operating fluid pressure in the supply tubing 16 and the exhaust ,pressure in the'return tubing 13 are 'applied to the engine section of the pump unit 24, under the control of an engine valve means 26, in a manner to produce reciprocatory movement of a combined engine and pump piston means 2S, Figs. 3 to 6, within the pump unit 24. The engine valve means 26 is incorporated in a free valve unit 30 which is hydraulically movable through the supply tubing 16 between the surface and an operating position in the housing 22 alongside of and spaced laterally from the pump unit 24, the pump and valve units thus being movable between their side-by-side operating positions4 and the surface independently of each other.

The supply, return and production tubings`16, 18 and 20 are connected at their upper ends to a control device 32 having therein valve means, not shown, operated by a handle 34 for connecting the supply, return and production tubings to supply, return and production lines 36, '38 and 4t) in various ways. The supply -line 36 is connected to a suitable source of operating tiuid under pressure, such as surface pumping equipment, not shown, while the return and production lines 38 and 4t) are connected to suitable points of disposal, not shown, for the spent operating fluid and the production iluid. When the pump and valve units 24 and 30 are in operation in the well, the handle 34 is moved to a position such as to cause the valve means in the control device 32 to connect the supply, return and production tubings 16, 18 and 20 to the supply, return and production lines 36, 38 and 40, respectively. However, under other operating conditions, such as removal nand/or installation of the pump unit 24 and/or the valve unit 30, the handle 34 may be moved to other-positions to cause the valve means within the control device-32 to interconnect the supply, return and production tubings 16, 18 and 20 and the supply, return and product-ion lines 36, 3S and 4t) in other ways, as required by the particular operation to be performed.

Surmounting the control device 32-and closing the uppernendsv of the supply and production tubings 16 and 20i are closures 42 and 44Vwhich are removable to permit insertion of the valve andl pump units' and 24 into the supply and production tubings 16A and 20,jrespectively, or' to permitremoval of the valve `and pumprunits from 1 and 2.0ithe drawings, the? the supply and production tubings. The closures 42 andl 44 may be provided with valve-unit and pump-unit catchers, respectively, for catching the valve and pump units as they are moved to the surface hydraulically from their operating positions within the housing 22.

The foregoing merely constitutes a general description of the structure and operation of the hydraulic pumping system of the invention, itbeing unnecessary to describe this system in more detail herein, except for certain details of the pump unit 24, since it is fully described in my aforementioned copending application. Accordingly, reference is hereby made to such copending application for a more complete disclosure of any subject matter not described in specific detail hereinafter.

Turning now to Fig. 3 of the drawings, the pump unit 24 is shown semidiagrammatically as including a housing 46 which provides a cylinder means 48 for the combined engine and pump piston means 28. The cylinder means 48 is provided at its lower end with an inlet 50 which communicates with the interior of the casing 10 at the level at which the pump housing 22 is set, the inlet being provided with a standing valve 52 lwhich prevents backow into the well. At the upper end of the cylinder means 48 is an outlet 54 which communicates y with the lower end of the production tubing 20 so that production fluid pumped from the well through the inlet- 50' and the outlet 54 is discharged into theproduction tubing for conveyance thereby to vthe surface. While the standing valve 52 is shown as located at the inlet 50,

it may be located at the outlet 54 also, as disclosed in my aforementioned copending application.

The combined engine and pump piston means 28 com-l prises a single plunger 56 having an intermediate -section of reduced diameter to provide upper and lower piston 35 or plunger elements 58 and 60 interconnected by a rod 62. The plunger 56 is provided with an axial production-fluid passage 64 through which the production iluid may flow from the inlet 50 to the outlet 54 along a straight, axial path, backflow through the passage 64 being prevented by a' working valve 66 carried by the plunger.

The plunger 56 is provided with an upwardly facing pump area 68, equal to the entire area of the upper end of the upper piston element 58, which supports the entire column ofA production uid in the production tubing 20 during the upward or working stroke ofthe plunger, v and 'which moves the production column upwardly" through the production tubing during such stroke. As will be apparent, the working valve 66 is closed'during the upward, working stroke of the plunger 56 and the standing valve 52 is open to permit flow of production uid from the well into the lower end of the cylinder means 48 through the inlet 5t) during such stroke. During the downward or return stroke of the plunger 56, the standing valve 52 is closed to prevent backow into the well and the working valve 66 is open to permit displacement of fluid from below the plunger 56 through the axial passage 64 into the cylinder means 48 above the plunger. Thus, the pump area 68 of the plunger 56 and the standing and working valves 52 and 66 constitute the pump section of the pump unit 24. v v

The engine section of the pump unit 24 includes a downwardlyfacing annular area 70 which is formed by the lower end of the upper piston element 58 and which' ating tiuid pressure in a cylinder 72 in which the piston element S8 reciprocates, the cylinder 72 being in constant communication with the supply tubing 16 through a v port 74 and the engine valve means 26 in a manner more fully `described in the aforementioned copending application. The engine section of the pump unit 24 also includes an upwardly-facing annular area 76 which is formed by the upper end of the lower piston element 60 and which encircles the rod 62'. The annular area 76 is' exposed to fluid pressure in a cylinder 78 in which the lower piston element 60 `is reciprocable, the cylinder 78 being separated fromy the cylinder 72 by a partition 80 having an axial bore 82 therethrough for the rod 62. The cylinder 78 is provided with a port 84 which is alternately connected to the supply tubing 16 and the return tubing 18 by the engine valve means 26 in a manner more fully described in the aforementioned copending application. subjected to the operating fluid pressure inf the supply tubing 16 and the spent operating'uid orV exhaust-pressure in the return tubing 18. i

With the foregoing construction, it will be apparent that when the area 76 is exposed tothe exhaust pressure, the operating lluid pressure constantly acting on the engine area 70 produces the working stroke of the plunger 56. The annular area 76 is slightly larger than, the engine area 70 so that, when the operating iluid pressure is applied to the annular area 76, the returnrstroke of the plunger 56 is effected despite the fact that the operating fluid pressure is constantly applied to the engine area '70. The necessary diiference between the engine area 70 and the annular area 76 may be provided by making the lower piston element 60 a few thousandths of an inch larger than the piston element 58.

It will be understood that instead of constantly applying the operating tiuid pressure to the area 70 and alternately applying the operating uid pressure and the exhaust pressure to the area '76, the working and return strokes may be eiected by alternately applying the operating tiuid pressure and the exhaust pressure to the area 70 while constantly applying the exhaust pressure to the area 76.

In the pump unit24 shown in Fig. 3 of the drawings, the ratio of the pump area 68 to the engine area 70 is only 1.065 :l and, as will be apparent, such a low pump! area to engine area ratio necessitates making the axial passage 64 of small diameter, and necessitates a correspondingly small working valve. 66, the standing valve 52 also being correspondingly small since there would be no point in making it mucli larger than the working valve. Fig. 3 is of small diameter even though the wall thickness of the tubular rod 62 has been reduced to an absolute minimum.

In Figs. 4, S'and 6 ofthe drawings, the pump unit 24 is shown as having pump area to, engine area ratios of 2:1, 3:1' and 51:1, respectivelythe same reference numerals being utilized throughout Figs. 3 to 6' in view of the correspondence of parts. Progressively increasing the ratio of the pump area 68 to the engine area 70 from l.065:1 to 5:1 results in progressively increasing the Consequently, the area 7.6 is alternately.

It will be noted that the axial passage 64 in outside diameter of the tubular connecting rod 62. This 1 provides for a progressive increase in the diameter of the axial production-Fluid passage 64 through the plunger S6, and corresponding progressive increases in the sizes of the standing valve 52 and the working valve 66. Also, the Wall thickness of the tubular connecting rod 62 is shown asprogressively increasing with the progressively increasing ratio of the pump area 68 to the engine area 7) to provide progressively improved structural characteristics. i

As will be apparent, increasing thepurnp area to engine area ratio greatly reduces the fluid friction losses through the axial production-fluid passage 64 because of the increased diameterof this passage and because of the increased sizes of the standing valve 52 and the working valve 66. With the 1.065:1 ratio of Fig. 3, the outside diameter of the plunger rod 62 is only 25% of the outside diameter of the piston element 5S of the plunger 56. Under these conditions, the diameter of the axial passage 64 `is necessarily so small that the production-Huid velocity therethrough must attain a value of more than twenty times the velocity of the-plunger 56. This results in serious ,-uid. friction losses, and presents a serious limitation on plunger .speed and pump capacity. By going tothe 2:1 ratio of Fig. 4, a substantial improvement results, the tubular plunger rod 62 in this case having a diameter equal to 70.7% of the diameter of the plunger 56. Under these conditions, the production-huid velocity through the axial passage 64 is only approximately four timesthe plunger velocity, with a wall thickness for the rod 62 adequate to provide the necessary structural strength.

However, optimum results are achieved by utilizing a ratio of the pumpiarea 68 to the engine area 70 of '3: 1, or higher, as shown in Figs. 5 and 6. With such ratios, the annular engine area 70 if reduced in size to such asmall proportion of the total cross-sectional area of the piston element 58 of the plunger 56 that satisfactorily low production-duid velocities through the axial passage 6,4, relative to plunger velocity, are achieved. Also, the. working valve 66 has the best proportions with a ratio of 3:1,V or higher, and the iluid velocities through the working valve seat and in the annulus between the working valve and thecylindercan be made approximatelyk equal, which is desirable. These factors all contribute. to reducing uid friction losses through the plunger 56 to a minimum with a pump area toengine area of 3:1 or higher.

Referring to Fig. 7 of the drawings, the graph presented therein illustrates the effect of increasing the ratio of the pump area 68 to .the engine area 70 on the ratio K. of the diameter of the tubular plunger rod 62 to the diameter ofthe piston element 58 of the plunger 56, the latter ratio being expressed in percentages. As Fig. 74

`clearly shows, the outside diameter of the tubular..

plunger rod 62,.and thus the diameter of the axial passage 64 and the sizes ofthe standing and working valves 52 and 66, increase rapidly with an increase in the v ratio of the pump area 68 to the engine area 70 until a ratio of 3:1 is attained, the ratio of plunger rod diameter to plunger diameter increasing slowly with an increase in the ratio of the pump area to the engine area abovev 3:1. In fact, the curve of Fig. 7 becomes so nearly ilat by the time aratio of the pump area 68 to the engine area 70 of 5:1 is achieved that, for all practical purposes, a ratio of 5:1 may be regarded as a maximum. It will be apparent that the side-by-side locations of the Vpump unit 24 and the valve unit 30 are, important inv achieving the best results with a ratio ofl the pump area 68 to the engine area 70 of 3: l, or more,lsiuce the engine valve unit does not interfere with the pump unit. This permits taking maximum advantage of the large production-duid passage 64 and the large working valve 56 permitted by the present invention since a correspondingly large standing valve 52 and correspondingly large pasg sages throughout the remainder of the pump unit 24 may be utilized.

Utilizing a pump area to engine area ratio of at least 3:1 requires aA correspondingly high operating iluid pressure to apply to the plunger 56 the same working-stroke i force as is attained with lower ratios and operatinguid pressures. As hereinbefore pointed out briefly, the super pressure operation of the present invention provides new and startling advantages and economies over any ,presently used hydraulic pumping system which may best be shown by comparing the results of super pressure operav tion with the result of operationat conventional pressures. The comparison will be-limited to oneset of4 operating conditions, it being. understood that similar comparisons maybe made for other sets of conditions.

Outlining the set of conditions under which-the comparison mentioned will be made hereinafter, it willrbe assumed that the pump unit 24 is a 21/2 inch unit, i.e., one which may be inserted into a 21/2 inch production tubing 20; vThe comparisonwill further be based on a@ pump depth setting of 10,000 feet and a production rate; Y of 350 barrels per day. Also, it will be assume'dthat`v the meansurfacetemperature is 60. F. andthatthe.,

f meneer" @D temperature gradient vinthe well is 1 F. for each 60 ft.'Y of depth, the temperature of course increasing with depth. p The production fluid is assumed to have a viscosity of 60 SSU at 100 F. and a specific gravity of 0.87.

Assuming the foregoing set of conditions, the fluid friction losses throughout the system were computed in terms of pressure drops in pounds per square inch. In computing the uid friction losses, each of the supply, return and production tubings l16, 18 and 20 was divided into sections of 1,000 feet in length and the fluid friction loss was computed for each such section of each tubing for the conditions obtaining at the midpoint thereof. The total iiuid friction loss in each of the tubings was then obtained by adding the fluid friction losses in each of the 10 sections into which it was divided.

Fluid friction loss computations were made in the foregoing manner for pump area to engine area ratios of 1:1, 2:51, 3:1, 4:1'an`d 5:1 for three different sets of tubing sizes, viz., a diameter for the production tubing 20 of 21/2 inches and diameters for the supply and return tubings 16 and 18 of 1 inch, a diameter for the production tubing 20 of 21/2 inches and diameters for the supply and return tubings of 3A inch, and a diameter for the production tubing of 21/2 inches and diameters for the supply and return tubings of 1/2 inch. For convenience, the sets of tubing sizes will be referred to hereinafter as the 21/2" X l system, the 21/2 X 3%: system and the 21/2" X 1/2" system, respectively. The comparative results for pump area to engine arearatios of 1:1, 2:1, 3:1, 4:1 and 5:1 for the 21/2 X l", the 21/2" X f and the 21/2" x 1/2 systems are tabulated in Tables I, II, III as follows:

Table I (Z1/2 x 1") Ratio 1:1 2:1 3:1 4:1 5:1

Friction in tubing 20Xratio 7 14 21 28 35 Friction in tubing 16- 434 211 201 235 206 riction in tubing 18. 356 112 63 40 30 Column density difieren 1 -70 -125 -170 -210 -245 Theoretical operating pres- Sure 3, 689 7, 380 1l, 067 14, 756 18, 445 Actual operating pressure. 4, 416 7, 592 11, 182 14, 849 18, 561 Net pressure loss. 727 212 115 93 116 Loss in percent of actua1 16. 4 2. 'i9 1. 03 626 625 Reciprocal of loss in percent. 0608 .358 972 1. 597 1.600

Table 1I (2l/2 x Ratio 1:1 2:1 3:1 4:1 5:1

Friction in tubin g 20Xratio- 1 7 14 21 28 35 Friction in tubing 16 l, 495 632 581 665 882 Friction in tubing 18 1,233 354 182 117 87 Column density differentiaL -105 -120 -170 -210 -250 Theoretical operating pressure 3, 689 7, 380 11,067 14, 756 18, 445 Actual operating pressure. 6, 320 8,260 11, 681 15, 356 19, 200 Net pressure loss 2, 631 880 614 600 755 Loss of percent of actual 41. 5 10. 66 5.25 3.91 3. 93 Reciprocal of loss in percent- 0240 0939 190 256 254 Table III (2l/2 x 16") Ratio 1:1 2:1 3:1 4:1 5:1

Friction 1n tubing 20 ratioA 7 14 21 28 35 Friction in tubing 16 9, 270 2, 900 2, 341 2, 959 3, 880 Friction in tubing 18 4, 931 l, 395 669 419 298 Column density differentialA -240 -175 -2051 -240 -280 Theoretical operating pressure 3,689 7,380 l1, 067 14, 756 18, 445 Actual. operating p essure 17, 657 11, 514 11, 893 17, 922 22, 378 Net pressure loss 13,968 4,134 2,826 3,166 3, 933 Loss in percent of actual-. 79.0 35.8 20.3 17. 65 17.55 Reciprocal of loss in percent. 0126 0279 0491 .0566 0569 in'the tubings specifiedin pounds per-square inch. It

should be pointed out that the friction loss i'n the tubing 20 is actually the addition to the operating .pressure as a result of 'this'frictionL *This operating pressureintuzevsev'ff includes the pump to engine area lratio factor. For'in: stance, the friction loss in the tubing 20, as shown in Tables I, II and III, is 7 p.s.i. gauge. With a 1:1 ratiothis is also a 7 p.s.i. gauge increase in the operating pressure, but with a 2:1 ratio, this becomes 14 p.s.i. gauge,1' and correspondingly more for the other ratios. The item Column density differential is a correction which takes into consideration the fact that the supply column in the tubing 16, being under higher pressure than the return column in the tubing 18, will have a higher density than the return column due to the compressibility of the oilA utilized as the operating fluid. The correction lrepre*l sented by the item Column density differential is ajj negative one which reduces the operating uid pressure f required at the lower end of the supply tubing 16. The'l item Theoretical operating pressure represents the Opfv erating lluid pressure in pounds per square inch necessary to produce the working stroke of the plunger 56 under the conditions hereinbefore specified n a frictionless system. The Actual operating pressure is equal to the Theoretical operating pressure plus the Friction in tubing 20 ratio, the Friction in tubing 16, the Friction in tubing 18 and the Column density differential. `Thel Net pressure loss in Tables I, II and III is -in each instance the difference between the Actual operating pressure and the Theoretical operating pressure. The item Loss in percent of actual is equal to the difference be'- tween the Actual operating pressure `and the,Theov retical operating pressure multiplied by and divided by the Actual operating pressure. The item Reciprocal of loss in percent is obtained by dividing one by the v Loss in percent of actual. l

It should also be pointed out that, in computing the results presented in Tables I, II and III, appropriate corrections were made for changes in friction due to viscosity reductions with temperature increases in the well bore with depth. Also, corrections were made. for viscosity changes with pressure based on the average pressures for the ten sections into which each of thetubings 16, 18 and 20 was divided for computation purposes.

The comparative results containedin Tables I, II and III are plotted in Figs. 8 and 9 of the drawings. Fig. 8 is a graph presenting the relation between the Loss in percent of actual and the ratio of the pump area 68 to the engine area 70 for the 21/2 X 1",'th"e 21/2 X and the 21/2" x 1/z" systems. Fig. 9 presents in graph form the relationship between the Reciprocal of loss in percent and the ratio of the pump area to the engine area for the three systems of Tables I, II and III.

As clearly shown by Tables I, II and III and by Fig. 8, the pressure loss due to lluid friction decreases rapidly i, as the ratio of the pump area 68 to the engine area 70 is increased. Considering Table I and the corresponding curve of Fig. 8 as an example, it will be seen that while the Net pressure loss is 16.4% with a ratio of 1:1,v the loss drops to 1.03% at a ratio of 3:1, to 0.626% at a ratio of 4:1 and to 0.625% at aratio of 5:1. Similar comparisons may be obtained from Tables I, II andIII and from Fig. 8.v Y

As Fig. 8 clearly shows, the friction losses for the 21/2" X l", the 21/2 X 3A" and the 21/2 X 1/2" systems reach minimum values at ratios of pump area to engine' f area of from 4:1 to 5 :1 and have substantially reached ;.I such minimum values at a ratio of 3:1. However, in- I- each instance, the pressure loss at a ratio of 2:1 is con siderably higher, and ata ratio of 1:1 is very much higherfl Consequently, it will be apparent that a ratio or .pump area to engine area of 3:1 represents a minimum for the purposes of the present invention. On the other hand,

a ratio of 5:1 represents a practical maximum because?-y each of the curves of Fig. 8 shows a minimum frictid ing to increase slightly above this ratio.

Fig. 9, wherein -the- -r" eciprocall of loss inperCent loss at approximately 5 :1, the net pressure losses'teird plotted versus the ratio of pump area to engine area, also illustrates thatva ratio of 3:1 is a minimum value and a ratio. of 5:1 is a practical maximum value. It will be noted that, in each instance, the curves of Fig. 9 show a curvature reversal at a ratio of slightly less than 3:1, thisreversal being the most pronounced in the curve for the. 21/2" X l system, but existing in the curves for the other two systems. also. The three curves virtually atten out between ratios of 4:1 and 5:1 and peak at a ratio of 5:1` in thesarne manner as the curves of Fig. 8, thereby establishing the ratio of 5:1 as a practical maximum.

` While the hereinbefore-presented comparison between the.16..4%4 net pressure loss with a 1:1 ratio and the 1.03% net pressure loss with a 3:1 ratio in the 21/2 x 1" system strikingly illustrates the advantages resulting from the present invention in drastically reducing fluid friction lossesy as the result of super pressure operation, an even more4 striking comparison can be obtained by considering thefeffect. of increasing the production rate from 350 barrels per day with this system to 1,000. barrels per day, all other conditions remaining constant. At a 1:1 ratio, the, net friction loss would be increased to approximately 631%., making the power transmission eiciency only 37%, as compared with a power transmission eiiiciency of 83.6% at a production. rate of 350 barrels per day. However, if a 3 :l ratio of pump area to engine area is utilized with a, production rate. of 1,000 barrels per day in the 2%," x l" system, then the net pressure loss due to friction is increased from 1.03%1to only approximately 7.8% forahighpower transmission eciency of 92%. Thus, itY will be seenA that the ratio or pump area to engine area 0f'3.: 1,. or higher, also permits attaining high power transmission eilciencies with high production rates, which is. an important feature of the. invention.

Another important comparison graphically illustrating the advantages of super pressure operation is `the power transmission capacity of a one inch tubing having a length of.l0,000 feet, Assuming losses not exceeding 10%, the horsepower capacity of one inch tubing for dilerent ratios of. the pump area 68 to the engine area 70 is as follows:

Table IV' natio-. 1:1. 211 3:1 4:1 5:1

Hnimmer... l 17.5 so 9i 14o 196 Another important advantage of the present invention is the essential saving in supply tubing and return tubing costs: As will be seen by comparing Tables I and Il, the supply and returntubings 16 and 1.8 may be reduced in= size from one inch to 1% inch with an increase in the netv pressure loss from 1.03% to only 5.25%, the increasesv in net pressure loss being even less at ratios of 4*.1 a-nd`5:l. Consequently, with pump area to engine area ratios of 3: l, or higher, the supply and return tubings lfand 18 may be reducedV in size from one inch to 3% inch'. with very little sacrifice in power transmission efl'ciency. Considering the dierence in cost of one inch and'fyt inchrtubing, utilizing 3A. inch tubing for the tubings 16 and 18 represents a saving of $1,200 for a 10,000 ft. pump depth. Further cost savings may be realized in this. respect by going to l/2 inch tubing. With 1/2 inch tubing, the net `pressure loss at a 3:1 ratio is still onlyv 20.3%` and is even less with ratios of 4:1 and 5:1. Making: this comparison in a little diierent manner, it will be .seen that by using a ratio of pump area -to engine area of. from 3:1 to 5:1 with 1/2 inch tubing for the supply and return tubings 16 and 18, the net pressure loss experienced is only slightly higher than that resulting from the use of one inch tubing at a 1:1 ratio (Table I). Thus, veryIv substantial cost savings with respect to the supply andfreturn tubings 16 and 18 are possible. with the present invention.

The smaller sizes. for the supply and return tubings 16.

and-1&whichthe, present invention permits also result in a cost` saving from another standpoint. For example.. the 21/2" x 1" system requires. a 7 inch casing, whereasv utilizing pump area to engine area ratios in the conventional rang'e. For example, a pump unit which is in serted in a 21/2 inch tubing may have a capacity as high as 1,200 barrels per day, as compared with the present maximum of 500. barrels per day. Table V, below, gives the pumping capacities; of pump units utilizing a 3:1 ratio at the piston speeds and fluid velocities which have been found to be practical.

Table V Pump size -inches l 1% 2 Zit 3 4 Plunger size .-do... is 1% 1% 25s 2% Capacity (b,ld,) i 300 600 l. 200 l, 900 3, 200

Another important advantage inherent in the super pressure operation of the present invention is the cushioning effect of the operating fluid. The bulk modulus. of compressibility of the crude oil utilized for the operating iluid is approximately 200,000 p.s.i. gauge. At 11,000 p.s.i. gauge, which) is approximately the operating iluidpressure at 10,000 ft. with a 3:1 ratio in the 21/2" x l" system, the operating fluid is compressed approximately Y 51/z%. `An air system operating at 285 p.s.i. gauge with a range of roperating pressure ofyl5 p.s.i. gauge would be the equivalent, which isja very moderate pressure for an air cushioning system. Thus, a hydraulic system which at low pressure is considered almost incompressible becomes a highly compressible or soft system at the extreme pressures contemplated by the present invention, which is an important feature.

As another illustration of the magnitude of this comi pressibility, the system discussed in the preceding paragraph requires the addition of 24.7 gallons of operating fluid thereto to bring the pressure from zero up to 171,000 p.s.i. gauge. Remembering that the general example hereinbefore discussed is based on a production rate of 350 barrels per day, the system therefore requires 117 barrels of operating fluid per day with `a 3:1 ratio of pump area to engine area. Consequently, 7.25 minutes of operation would be required to compress the operatingfluid column in the supply tubing 16 to the full operating pressure of 11,000 p.s.i. gauge. Also, if the pump unit 24 were operating at a normal rate of 7l strokes per minute and were then stopped for one complete stroke with the delivery of operating'uid to the supply tub-Y Another point of considerable importance is the fact that the extreme pressure operation of the present invention results in onlyl a very slight increase in leakage losses in the pump unit 24. The reason for such a slight increase in leakage losses stems from the fact that super.

pressure operation greatly increases the viscosity of the.v

operating fluid. For example, if the viscosity of the operating fluid at 10,000 feet at an operating uid' pressure of 4,416 p.s.i., which is the operating fluid pressure set forth in Table I for a 1:1 ratio of pump area to engine area, is 2.9 centistokes, then theviscosity at the.Y sameA depth and an operating fluid pressure of 11,182 `p.s.'i: gauge, corresponding to thei3z1 ratio ofA Table'l, is 5.2 centistokes. Thus, while the pressure of 11,182l p.s.i.

l gauge corresponding to: a 3:1 ratio4 tends to produce more leakage than the pressure of 4,416 p.s.i. gauge corresponding to a 1:1 ratio, this tendency is partially offset by the corresponding viscosity increase from 2.9 centistokes to 5.2 centistokes resulting from the higher pressure with a 3:1 ratio. It can be shown that the combined effect of the increased leakage due to the increased pressure and the decreased leakage due to the increased viscosity resulting from the increased pressure results in a leakage of only 1.7 units for a 3:1 ratio of pump area to engine area for each unit of leakage with a 1:1 ratio of pump area to engine area. In other words, going from a 1:1 ratio of pump area to engine area to a 3:1 ratio results only in a 1.7 :1 increase in leakage losses. Consequently, all of the hereinbefore-discussed advantages of super pressure operation land pump area to engine area ratios of at least 3:1 are attained with only a 1.7:1 increase in leakage in the pump unit 24.

Although various specific illustrative examples have been presented herein, it will be understood that the nvention is not to be regarded as limited thereto and is to be accorded the full scope of the following claims'.

I claim: i

l. ln a fluid operated pump, the combination of: cylinder means having upper and lower ends and having an inlet adjacent its lower end and an outlet adjacent its upper end; combined engine and pump piston means reciprocable in said cylinder means and having an axial passage therethrough for production iluid pumped by said pump which communicates at its lower end with said inlet and at its upper end with said outlet, said piston means having at the upper end thereof an upwardly-facing, transverse pump area and having below the upper end thereof a downwardly-facing, annular, transverse engine area which encircles and is spaced radially outwardly from said axial passage, the ratio of said pump area to said engine area being at least 3:1, a working valve carried by said piston means and controlling flow through said passage; a standing valve carried by said cylinder means and controlling ow through one of said inlet and said outlet; and means for applying operating fluid pressure to said engine area.

2. In a fluid operated pump, the combination of: cylinder means having upper and lower ends and having an inlet adjacent its lower end and an outlet adjacent its upper end; combined engine and pump piston means comprising a plunger reciprocable in said cylinder means and having an axial passage therethrough for production uid pumped by said pump which communicates with said inlet at its lower end and with said outlet at its upper end, said plunger having at its upper end an upwardlyfacing, transverse pump area and having intermediate its ends a section of reduced diameter providing a downwardly-facing annular, transverse engine area which encircles and is spaced radially outwardly from said axial passage and providing an upwardly-facing, annular, transverse engine area which also encircles and is spaced radially outwardly from said axial passage, the ratio of said pump area to said downwardly-facing engine area being at least 3:1, said upwardly-facing engine area being slightly larger than said downwardly-facing engine area; a working valve carried by said plunger and permitting upward ow through said axial passage; a standing valve carried by said cylinder means and permitting upward ow through one of said inlet and said outlet; means for constantly applying an operating fluid pressure to said downwardly-facing engine area; and means for alternately applying said operating fluid pressure and an exhaust pressure to said upwardly-facing engine area.

3. ln a fluid operated pump, the combination of: cylinder means having upper and lower ends and having an inlet adjacent its lower end and an outlet adjacent its upper end; combined engine and pump piston means comprising a plunger reciprocable in said cylinder means and having an axial passage therethrough for production ud pumped by said pump which communicates with said inlet a: is lower ene and with, said butler at its vupperend, said plunger having at its upperv end an upwardly.

facing, transverse pump area and having intermediate its ends a section of reduced diameter providinga' downwardly-facing annular, transverse engine area which encircles and is spaced radially outwardly from'said axial passage and providing an upwardly-facing, annular, trans-vv verse engine area which also v encircles and is spaced radially outwardly from said axial passage, the ratio of said pump area to said downwardly-facing engine area being at least 3:1, said upwardly-facing engine area being ow through one of saidinlet and said ouden-means'.

for. constantly applying an operating'v tluid pressure to said downwardly-facing engine area; and engine valvemeans located alongside of and spaced laterally "froml said plunger for alternately applying said operating fluidipressure and an exhaust pressure to said upwardly-facing engine area.

4. In a fluid operated pumping system, the combination of: production and supply tubings set in a well and extending downwardly therein from the surface to a pumping zone; uid operated pumping means connected to said production and supply tubings in said pumping zone, and provided with an inlet in communication with the well and an outlet in communication with said production tubing, for pumping production fluid from the well through said inlet and said outlet into said production tubing, said pumping means including engine and pump piston means reciprocable through working and return strokes and having a transverse pump area which communicates with said outlet and which supports a colum of production fluid extending from said pumping zone to the surface in said production tubing during said working stroke of said piston means, and said piston means having a transverse engine area which faces in the opposite direction from said pump area and which is capable of being acted on by operating fluid under pressure in said supply tubing to produce said working stroke of said piston means, the ratio of said pump area to said engine area being at least 3:1; and means for applying operating fluid pressure from said supply tubing to said engine area.

5. A fluid operated pumping system as set forth in claim 8 wherein said pumping means includes an axial passage for production Ailuid being pumped which extends through said piston means and which communicates with said inlet and said outlet, said piston means carrying a working valve which controls ilow through said axial passage and said pumping means including a standing valve controlling ow through one of said inlet and said outlet.

6. In a uid operated pump, the combination of: cylinder means having upper and lower ends and having an inlet adjacent its lower end and an outlet adjacent its upper end; combined engine and pump piston means comprising a plunger reciprocable in said cylinder means and having an axial passage therethrough for production iluid pumped by said pump which communicates with said inlet at its lower end and with said outlet at its upper end, said plunger having at its upper end an upwardly-facing, transverse pump area and having intermediate its ends a section of reduced diameter providing a downwardly facing annular, transverse engine'area which encircles and is spaced radially outwardly from said axial passage and providing an upwardlyfacing, annular, transverse engine area which also encircles and is spaced radially outwardly from said axial passage, the ratio of said pump area to said downwardly-facing engine area being at least 3:1, said upwardly-facing engine area being slightly larger than said downwardly-facing engine area; a working valve carried by said plunger and petrmittingupward 'flow throughsaid axial passage; a stand-- ing valve carried by said cylinder means and permitting upward flow through one of said inlet and said outlet; means for constantly applying an exhaust pressure to said upwardly-facing engine area; and means for alternately applying said exhaust pressure and an operating fluid pressure to said downwardly-facing engine area.

1 7. In a uid operated pump, the combination of: cylinder means having upper and lower ends and having an inlet adjacent its lower end and an outlet adjacent its upper end; combined engine and pump piston means cornprising a plunger reciprocable in said cylinder means andhaving au axial passage therethrough for production fluid pumped by said pump which communicates with said inlet` at its lower end and with said outlet at its upper end, said plunger having at its upper end an upwardly-facing, transverse pump area and having intermediate its ends a section of reduced diameter .providing a downwardly-facing annular, transverse engine area which encircles and is spaced radially outwardly from said axial passage and providing an upwardly-facing, annular, transverse engine area which also encircles and is spaced radial-,-

1y outwardly -from said axial passage, the ratio of said pump area tosaid `downwardly-facing engine area being at 4least 3:1, said upwardly-facing engine area being slightly larger than said do.vnwardly-facing engine area; a working valve carried by said plunger and permitting upward dow through said axial passage; a standing valve carried by said cylinder means and permitting upward flow through one of said inlet and said outlet; means for con stantly applying an exhaust pressure to said upwardlyfacing engine area; and engine valve means located alongside of and spaced laterally from said plunger for alternately applying said exhaust pressure and an operating Huid pressure to said downwardly-facing engine area.

References Cited in the le of this patent UNITED STATES PATENTS 825,950 Weir s July 17, 1906 1,765,457 Shutt June 24, 1930 2,018,215 Lausen Oct. 22, 1935 2,127,168 Grant Aug. 16, 1938 2,132,084 Reiss Oct. 4, 1938 2,747,511 Turner et al May 29, 1956 2,751,14r Troendle lune 19, 1956 2,780,171 Heddy Feb. 5, 1957 UNITED STES P'INT OFFICE CERTIFCTE 'l RECTION Patent No., 2,329,327 Match 22 1960 Clarence Je Coberly It is hereby certified that error appears in the printed specification of the above numbered patent Tequrng correction and that the said Letters Patent should read as corrected below.

Column 2q line $39(l for "comprise" read -m compromise m15 column 9U- Table lli first column thereofv. next to the last lnez for "Loss of percent of actual" read n Lose in percent of actual fm; column 12,1 Table V(l first column thereof last lne for UDV/d J read (lo/d) @-55 column 13Il line 36(I for "3:11-'3' read m- 321; fe.

Signed and sealed this 30th day of August 1960.,

(SEAL) Attest:

ERNEST We SWIDER RoBERT c. WATSON Attesting Officer Commissioner of Patents 

